journal-bearing-vibration-and-ssv-hash

The test bearings consist of a conventional center pivot bearing with labyrinth end seals Figure 8 and evacuated direct lube center and offset pivot bearings.. Although there were more i

Scan M DeCamillo is Manager of Research and Development for Kingsbury, Inc., in Philadelphia, Pennsylvania He is responsible for design, analysis, and development of Kingsbury fluid film bearings for worldwide industrial and military applications He began work in this field in 1975 and has since provided engineering support to industry regarding application and performance of hydrodynamic

bearings Mr DeCamillo has developed performance and

structural bearing analysis tools during his career, establishing

design criteria used in many publications and specifications.

He has patents and has authored several papers on

bearing research, which is currently focused on advancing

hydrodynamic bearing technology in high-speed turbomachinery.

Mr DeCamillo received his B.S degree (Mechanical

Engineering, 1975) from Drexel University He is a registered

Professional Engineer in the State of Pennsylvania and a member

of STLE, ASME, and the Vibration Institute.

Minhui He is a Machinery Specialist with BRG Machinery Consulting LLC, in Charlottesville, Virginia His responsibilities include vibration troubleshooting, rotor-dynamic analysis, as well as bearing and seal analysis and design He is a member of STLE, and is also conducting research on rotordynamics and hydrodynamic bearings.

Dr He received his B.S degree (Chemical Machinery Engineering, 1994) from Sichuan University From 1996 to 2003, he

conducted research on fluid film journal bearings in

the ROMAC Laboratories at the University of Virginia,

receiving his Ph D (Mechanical and Aerospace

Engineering, 2003).

C Hunter Cloud is President of BRG Machinery Consulting, LLC, in Charlottesville, Virginia, a company providing a full range of rotating machinery technical services He began his career with Mobil Research and Development Corporation in Princeton, NJ,

as a turbomachinery specialist responsible for application engineering, commissioning, and troubleshooting for production, refining, and chemical facilities During his 11 years at Mobil, he worked on numerous projects, including several offshore gas injection platforms in Nigeria, as well as serving as reliability manager at a large US refinery.

Dr Cloud received his B.S (Mechanical Engineering, 1991) and Ph.D (Mechanical and Aerospace Engineering, 2007) from the University of Virginia He is a member of ASME, the Vibration Institute, and the API 684 rotordynamics task force.

James M Byrne is currently a member of the BRG Machinery Consulting team, in Charlottesville, Virginia BRG performs research and analysis in the fields of fluid film bearings, magnetic bearings, and rotordynamics Mr Byrne began his career designing internally geared centrifugal compressors for Carrier in Syracuse, New York He continued his career at Pratt and Whitney aircraft engines and became a technical leader for rotordynamics Later Mr Byrne became a program manager for Pratt and Whitney Power Systems managing the development of new gas turbine products From 2001 to 2007,

he was President of Rotating Machinery Technology, a manufacturer

of tilting pad bearings.

Mr Byrne holds a BSME degree from Syracuse University, an MSME degree from the University of Virginia, and an MBA from Carnegie Mellon University.

JOURNAL BEARING VIBRATION AND SSV HASH

by Scan M DeCamillo Manager, Research and Development

Kingsbury, Inc.

Philadelphia, Pennsylvania

Minhui He Machinery Specialist

C Hunter Cloud President and James M Byrne Machinery Consultant BRG Machinery Consulting, LLC Charlottesville, Virginia 22903 USA

11

Peculiar, low-frequency, radial vibrations have been observed in

various turbomachinery using tilt-pad journal bearings Unlike

discrete subsynchronous spikes that often indicate a serious

problem, the vibrations are indiscrete and of low frequency and

amplitude The low level shaft indications have raised concern in

witness tests of critical machinery, even in cases that comply with

American Petroleum Institute (API) limits, owing to uncertainty

regarding the cause and nature of the vibrations

This paper presents shaft and pad vibration data from various

tilt-pad journal bearing tests that were undertaken to investigate and

better understand these subsynchronous indications The vibration

characteristics are defined and compared under the influence of

speed, load, oil flow, and bearing orientation Results are presented

for conventional and direct lube tilt-pad bearing designs, along with

discussions of parameters and methods that were successful in

reducing and eliminating these low level vibrations

The test results indicate that the low-frequency shaft indications

are caused by pad vibration Hypotheses and analyses are

presented and discussed in relation to the test observations

INTRODUCTION

Many early technical papers report on the inherent stability of

tilt-pad journal bearings in overcoming oil whirl and oil whip

limitations of fixed geometry bearings There are, however, other

vibration phenomena associated with tilt-pad bearings that are

topics of past and present research In the late 70s, extensive

babbitt fatigue cracking of upper, unloaded pads was a major

problem in large, conventional tilt-pad journal bearings These

were flooded designs, which incorporate end seals to restrict oil

outlet and flood the bearing cavity A well-referenced study by

Adams and Payandeh (1983) attributes the damage to self-excited,

subsynchronous pad vibration

Trends for larger and higher speed turbomachinery impose

greater demands on bearings and rotor dynamics Direct lube

bearings have evolved to reduce the higher power losses, oil flow

requirements, and pad temperatures associated with higher surface

speeds Direct lubrication is not a new concept as designs have been

used in special applications with a long history of reliability Use of

direct lube journal bearings became more prevalent as machine size

and speed increased, eventually spawning several papers in the early

90s investigating their steady-state performance These include

research by Booser (1990), Tanaka (1991), Harongozo, et al

(1991), Brockwell, et al (1992), and Fillon, et al (1993)

The references document pad temperature reductions derived

from the efficient evacuation of hot oil from the bearing cavity

Direct lube bearings are therefore typically designed for evacuated

operation, accomplished by opening up end seal and oil outlet

restrictions A direct application of oil to the journal surface

prevents oil from bypassing the films, which can occur in

conventional bearings when outlet restrictions are removed Use of

nozzles to spray oil on the journal surface between pads is one

method of direct lubrication An additional pad temperature

advantage is gained by direct lubrication features

Literature on vibration characteristics of direct lube journal

bearings is not as prevalent DeCamillo and Clayton (1997) present

rotordynamic data for large, 18 inch (457 mm) generator bearings

The tests showed comparable vibration response for conventional

and direct lube designs Edney, et al (1996), provide similar

information for a small, high-speed, multistage steam turbine with

4 inch (102 mm) diameter journal bearings Peculiar, low-frequency,

radial vibrations were observed during these steam turbine tests, but

levels were low and acceptable However, similar vibrations in a

high-speed compressor during acceptance tests in 1999 did encroach

upon acceptable API limits, and significant time and resources were

expended to address this issue (DeCamillo, 2006)

Personal experience and discussions among original equipment

manufacturers (OEMs) and users over the past few years indicate

that these low-frequency vibrations have been encountered in turbines, compressors, and gearboxes, using conventional and direct lube tilt-pad bearing designs The signature has also been documented in separate research investigating stability (Cloud, 2007) Accurate stability prediction is a major topic of concern in the industry Kocur, et al (2007), highlight a large spread in predicted bearing stiffness and damping coefficients among computer codes, and associated ramifications regarding stability assessment of critical turbomachinery At the same time, there are not many codes available that address direct lube performance and dynamic coefficients (He, 2003; He, et al., 2005)

Researchers are presently attempting to sort through some difficult questions What is the source of the low-frequency vibrations? Are they attributable to direct lube bearings, pad flutter, starvation, high speeds, or low loads? Do they affect machine stability, safety,

or reliability?

This paper presents a chronology of tests, investigations, and theoretical analyses aimed at providing answers to some of these questions It is desired that the information be of value to researchers, OEMs, users, and other personnel involved with hydrodynamic bearings and vibration in turbomachinery

SSV HASH, DEFINITION Figure 1 is an example radial vibration spectrum from a high-speed compressor test using direct lube tilt-pad journal bearings The main subject of this paper is the low-frequency shaft vibrations indicated in the figure Unlike discrete subsynchronous spikes that often indicate a serious problem, the vibrations are indiscrete and

of low-frequency and amplitude The plot is paused to capture the random, broadband frequencies In order to distinguish the subsynchronous vibration (SSV) characteristics under consideration,

the term SSV hash is defined for use in this paper:

• SSV Hash: A vibration signature characterized by low-frequency, low amplitude, broadband subsynchronous vibrations that fluctuate randomly.

Figure 1 Example Radial, Low-Frequency, Broadband, Vibration.

Experience with turbomachinery bearing sizes 3.88 to 8.00 inches (100 to 200 mm) in diameter has observed SSV hash frequencies typically ranging up to 30 Hz with amplitudes on the order of 0.1 to 0.2 mils (.0025 to 0050 mm) peak-to-peak

INITIAL INVESTIGATIONS Delayed acceptance of a high-speed compressor due to SSV hash

in 1999 prompted a research project performed on a high-speed test rig described in detail in a separate reference (Wilkes, et al., 2000) Pertinent information is provided here for convenience The rig has

a test shaft driven by a variable speed gas turbine through a flexible coupling (Figure 2) The test shaft is approximately 5 feet (1.5 m) long and 5 inches (127 mm) in diameter, supported at either end by

a pivoted shoe journal bearing Two orthogonal proximity probes are mounted inboard of each journal bearing, ± 45 degrees off top-dead-center, to record radial shaft vibration A spectrum analyzer was used to acquire fast Fourier transform (FFT) vibration

signatures, presented in paused plots in Figures 3 through 5 The

y-axis represents peak-to-peak amplitude in mils

Figure 2 High-Speed Test Rig Journal Bearings and Shaft Schematic.

Figure 3 Evacuated Discharge Configuration, 3.0 gpm (11.4 l/min).

Figure 4 Flooded Discharge Configuration, 6.0 gpm (22.8 l/min).

Figure 5 Evacuated Discharge Configuration, 3.0 gpm (11.4 l/min),

SSV Grooves.

Fortunately, it was possible to duplicate the low-frequency vibration signature in the test rig, which allowed a parametric study of many journal bearing designs and configurations over the course of this initial investigation The data presented in Figures 3 through 5 are for direct lube, five-pad, leading-edge-groove journal bearing tests at a shaft speed of 10,000 rpm and a low, 20 psi (0.14 MPa) projected load The bearings have a nominal diameter of 5 inches (127 mm) and an axial length of 2.25 inches (57 mm) The pads are steel backed with a babbitt surface and have a 60 degree angle and a 60 percent offset pivot The assembled bearing diametric clearance is 0.009 inch (0.23 mm) and the nominal preload is 0.15 Tests were run with ISO VG 32 turbine oil supplied at 120⬚F (49⬚C)

The SSV hash indicated in Figure 3 was recorded for an oil flow of 3.0 gpm (11.4 l/min) for the direct lube bearing in its as-designed, evacuated oil discharge configuration Tests found that increasing the oil flow tended to reduce the amplitudes but did not entirely eliminate the SSV hash signature Elimination required the installation of floating oil seals as well as an increase

in oil flow to 6.0 gpm (22.8 l/min), the results of which are shown

in Figure 4 This solution unfortunately required higher oil flow and operated with higher power loss and pad temperatures than the original design

Methods were therefore pursued to address these performance issues, and results began to suggest that the low frequencies may be due to air entering the oil film Based on this hypothesis,

a design was conceived that might eliminate SSV hash while maintaining some direct lube benefits The pads were modified

with narrow circumferential SSV grooves, cut in the babbitt

near the edges of the pads (Figure 5), to capture and redirect side leakage toward the leading edge of the next pad (Wilkes and DeCamillo, 2002) In this way, additional oil is made available to the oil films without increasing the bulk oil flow to the bearing

The SSV groove pads were installed and tested in the original evacuated condition (oil seals removed) with results shown in Figure 5 for 3.0 gpm (11.4 l/min), and tested as low as 2.0 gpm (7.6 l/min) with negligible SSV hash indications Of several methods pursued during the course of this initial investigation, the grooves were the only solution successful in eliminating SSV hash in an evacuated configuration The low oil flow and power loss of the original design were maintained, with a slight penalty in pad temperature due to the introduction of warm, side leakage oil back into the oil film

Another observation of this initial investigation was the tendency for an increase in synchronous amplitudes when SSV hash was eliminated, noticeable by comparing both the flooded solution (Figure 4) and the evacuated groove solution (Figure 5) with the original open discharge configuration data

of Figure 3

SUBSEQUENT TESTS Although solutions were obtained in initial investigations, operating conditions were limited in load and speed A subsequent series of tests was initiated in 2005 to further investigate the low-frequency vibrations The same test rig was modified to incorporate a radial load system in the bearing housing at the free end of the shaft, shown schematically in Figure 6 The system incorporates a hydraulic cylinder that loads through a dowel to a single journal bearing loader shoe on top of the shaft This pushes down on the shaft and loads the test bearing The applied load is measured by a load cell below the hydraulic cylinder A new data acquisition system was used to acquire the data, and additional proximity probes were installed to monitor the vibration of an upper and a lower journal bearing pad The pad proximity probes were mounted in the bearing casing and targeted the trailing edge

of the back of the pads at the locations depicted schematically

in Figure 7

13

JOURNAL BEARING VIBRATION AND SSV HASH

Figure 6 Test Rig Radial Load System Schematic, Free End of Shaft.

Figure 7 Bearing Orientation and Pad Numbering.

The test bearings consist of a conventional center pivot bearing

with labyrinth end seals (Figure 8) and evacuated direct lube center

and offset pivot bearings Three direct lube methods were tested:

spray nozzles (Figure 9), leading-edge-grooves (Figure 10), and

between-pad-grooves (Figure 11) Each bearing was tested for

load-on-pad and load-between-pad orientation over a range of

loads, oil flows, and speeds

Figure 8 Conventional Bearing with Labyrinth End Seals.

Figure 9 Direct Lube Spray Nozzles, Evacuated Configuration.

Figure 10 Direct Lube Leading-Edge-Groove, Evacuated Configuration.

Figure 11 Direct Lube Between-Pad-Groove, Evacuated Configuration.

Bearing geometry, test oil viscosity, and inlet temperature are the same as in earlier investigations except that the assembled diametric clearance is 0.0072 inch (0.18 mm) and the nominal preload is 0.25 Loads indicated in the following sections and figures refer to the applied radial load For data labeled no-load, the journal bearings are supporting the weight of the shaft, which gives a 20 psi (0.14 MPa) projected unit bearing load Each test bearing is installed around the free end of the shaft without disturbing the coupling or the coupling end journal bearing The same coupling end journal bearing is used in all tests Vibration data were recorded while varying speed to 14,000 rpm, load to 400 psi (2.8 MPa), and oil flow to 30 gpm (114 l/min), holding two of the parameters constant while the other was varied

The speed ramp data are used in this paper, presented in color-coded FFT waterfall diagrams with a vertical scale denoting peak-to-peak amplitude in mils With Figure 12 as an example, each waterfall diagram contains data for three separate run-up/run-down speed ramps from 5000 to 14,000 to 5000 rpm (83 to 233 to 83 Hz), one for each of three oil flows The back plane of the diagram contains a projection of all data Vibrations below 1 Hz are cut off to delete spurious 0 Hz indications from obscuring the low frequencies under investigation and the data are uncompensated, the main focus being the low-frequency vibrations As a guide, SSV hash amplitudes above 0.1 mils (0.0025 mm) peak-to-peak are levels that have caused concern

in acceptance tests

Figure 12 Conventional Center Pivot, LBP, No-Load.

Shaft SSV Hash Trends

The volume of data for the different bearing designs allowed for an

assessment of trends in shaft SSV hash This section describes those

that were fairly common for all test bearings Conventional bearing

speed ramp data in Figures 12 through 15 are used for reference

Figure 13 Conventional Center Pivot, LOP, No-Load.

Figure 14 Conventional Center Pivot, LBP, 400 psi (2.8 MPa).

Figure 15 Conventional Center Pivot, LOP, 400 psi (2.8 MPa).

Comparing data among the figures, it can be noticed that most SSV hash indications occur at low flow and low load The word

“indications” is carefully used in this generalization Although there were more indications at low flow and low load for most

bearing designs, SSV hash amplitudes were sometimes higher at

other operating conditions, as evidenced in results later in this paper Other common trends derived from the study are that SSV hash levels for no-load data were similar for load-between-pad (LBP) and load-on-pad (LOP) orientations (e.g., Figures 12 and 13) A curious result was that while SSV hash decreased with applied load for load-between-pad orientation (e.g., Figure 12 versus 14) there was only a small change with load for load-on-pad orientation (e.g., Figure 13 versus 15) Consequently, load-on-pad orientation had higher SSV hash levels than load-between-pad orientation at higher loads for all test bearings

It is worth noting that initial investigations and field reports were for high-speed applications with relatively light rotors, and so SSV hash was formerly associated with high-speed and low-load The test data in these subsequent series of tests indicate that this is not exactly true Although indications were more pronounced at low loads, tests indicate there can be undesirable levels of SSV hash in the case of load-on-pad orientation at higher loads, for example Figure 15 at 400 psi (2.8 MPa) There were also many indications

at lower speeds in the test data, again using Figure 15 as an example

Conventional Versus Direct Lubrication

Figure 16 displays no-load data for the conventional center pivot test bearing with load-on-pad orientation, the same used in Figure

13, but zoomed in on the lower frequencies and amplitudes under investigation Figure 17 contains comparable, center pivot, between-pad-groove direct lube data A condition mentioned earlier is noticeable in the direct lube data, i.e., there are more SSV hash indications at low flow but with higher amplitudes at an intermediate condition, 8 gpm (30 l/min) and 11,000 rpm in this particular test The figure also supports observations from field experience and initial investigations where SSV hash decreases, but is not necessarily eliminated at higher flows

Comparisons between conventional and direct lube results (Figures

16 and 17) are fairly straightforward Direct lube data have noticeable SSV hash indications at 8 gpm (30 l/min) whereas conventional bearing levels are negligible Conversely, very high subsynchronous vibrations are noticeable in conventional bearing data at 4 gpm (15 l/min) where the direct lube design has significantly lower SSV hash indications

These data indicate that flooding and oil flow are not the only key parameters influencing SSV hash Conventional bearing tests with labyrinth seals (Figure 16), and initial tests with oil seal rings

15

JOURNAL BEARING VIBRATION AND SSV HASH

(Figure 4), both required some increase in oil flow to eliminate the

signature, whereas higher oil flow is not as effective in evacuated

designs (e.g., Figure 17) This suggests cavity pressure may be

another influential parameter that, unfortunately, was not measured

Calculated pressure drops across oil discharge restrictions indicate

that cavity pressures between 0.5 and 2.0 psig (0.04 and 0.14 bar)

were present when SSV hash was noticeably reduced Also fitting

this scenario are the lower SSV hash indications in direct lube

designs compared to the conventional bearing data at low flow

(Figure 17 versus 16), which may be attributable to the more direct

application of oil to the oil films

Figure 16 Conventional Center Pivot, LOP, No-Load.

Figure 17 Direct Lube Between-Pad-Groove, Center Pivot, LOP,

No-Load.

Center Versus Offset Pivot

Center versus offset pivot tests were conducted for several direct

lube bearing configurations Comparisons have been difficult to

quantify For example, spray nozzle offset pivot data (Figure 18)

have high amplitudes and a broad band of SSV hash midway through

the 4 gpm (15 l/min) data speed ramps at approximately 12,000

rpm, indicated in the figure At the same conditions, spray nozzle

center pivot data (Figure 19) have less broadband indications, and

amplitudes are actually highest at approximately 6000 rpm In 8 and

12 gpm data (30 and 45 l/min), the offset pivot has slightly less SSV

hash Similar comparisons for between-pad-groove tests (Figure 17

versus 20) show lower SSV hash levels in the offset pivot data at

4 gpm (15 l/min), and only subtle differences at the higher flows

Figure 18 Direct Lube Spray, Offset Pivot, LOP, No-Load.

Figure 19 Direct Lube Spray, Center Pivot, LOP, No-Load.

Figure 20 Direct Lube Between-Pad-Groove, Offset Pivot, LOP, No-Load.

The results were unexpected in that offset pivot designs require more oil flow to the films, and it was suspected that this would be clearly reflected in SSV hash levels The data indicate there are factors other than bulk oil flow that need to be taken into account regarding SSV hash and pivot offset

Variations Among Direct Lube Designs

Three offset pivot, direct lube bearings were tested to compare

shaft SSV hash characteristics among different methods of direct

lubrication All three have an evacuated discharge configuration

The methods include spray nozzles, leading-edge-grooves, and

between-pad-grooves Photographs of the direct lube features are

provided in Figures 9 through 11, and corresponding vibration data

are contained in Figures 19 through 21

Figure 21 Direct Lube Leading-Edge-Groove, Offset Pivot, LOP,

No-Load.

In general, all three direct lube methods display noticeable SSV

hash indications Variations in amplitudes and frequencies make it

difficult to generalize the results The spray nozzles produced the

highest levels of SSV hash The leading-edge-groove operated

with lower SSV hash amplitudes but at higher frequencies, and

between-pad-groove SSV hash levels were lower for all conditionsin

comparison with spray nozzle data Results comparing

between-pad-groove and leading-edge-between-pad-groove data vary with operating conditions

SSV hash levels at 4 gpm (15 l/min) are negligible and lower than the

leading-edge-groove At higher flows, between-pad-groove SSV hash

amplitudes are higher but occur at a lower frequency

Distinct differences in shaft SSV hash signatures are obvious

among the designs Some speed dependence is noticeable in spray

nozzle data (Figure 18), and more so in leading-edge-groove data

(Figure 21), which appears to be suppressed at higher speeds

Between-pad-groove characteristics (Figure 20) do not show

variations with speed, and seem sensitive to a particular speed band

depending on flow There are also many conditions throughout

Figures 18 through 21 where the evacuated direct lube bearings

operate with negligible SSV hash indications

While the reasons for the variations are still under investigation,

the fact that there are differences provides valuable information

Since the pad geometry is the same for the different bearings, the

direct lube feature itself is influencing the shaft SSV hash and,

because these are evacuated designs, the influence is most likely

occurring at the entrance or leading edge of the oil film

Pad/Shaft Correlation Investigation

A review of 2005 vibration data from the upper and lower pad

proximity probes (depicted in Figure 7), noted more subsynchronous

indications in the upper pad, with the highest amplitudes occurring at

high-load and low-flow conditions in many cases This is reasonable

considering that the upper pad clearance increases as the shaft is

pushed down under load, away from the upper pad The upper pad oil

film consequently generates lower hydrodynamic forces and requires

more oil flow to fill the gap Unfortunately, an unexpected result of

the investigation was that there were only a few random correlations

between upper pad and shaft SSV hash for the broad range of

operating conditions and bearing configurations tested

This lack of correlation prompted another series of tests in

2006 with all pads monitored by proximity probes, installed and positioned as described earlier, and made possible by the acquisition of new high-speed vibration equipment Figure A-1 (refer to APPENDIX A) organizes shaft and pad vibration data for a leading-edge-groove bearing test at 200 psi (1.38 MPa), for a load-on-pad configuration as an example The waterfall diagrams are arranged so that data for the orthogonal shaft probes and all five pads can be viewed in relation to one another The same scales are used for all waterfall diagrams in the figure It is important to note that the shaft was removed and refurbished between 2005 and 2006 tests and a different coupling end journal bearing installed, so data are not comparable

to information from earlier tests

Referring to Figure A-1, it can be noticed that the subsynchronous vibrations in unloaded pad 2 at 4 gpm (15 l/min) do not correlate with either of the orthogonal shaft probes (labeled +45 degrees and

⫺45 degrees in the figure) A unique observation is that the side pads 3 and 5, which were not monitored in previous tests, have the strongest subsynchronous vibrations and show excellent correlation with the shaft SSV hash indications

Pad/Shaft Correlation General Trends

The observations from Figure A-1 relate to a specific bearing and set of operating conditions Comparisons of data for all five pads, over a broad range of operating conditions for the different bearing designs allowed for a more precise assessment of pad/shaft interaction The following observations and trends were determined

to be fairly common for all test bearings

All shaft SSV hash indications observed in these tests were confirmed to correlate with vibrations from at least one of the five pads The converse is not true There were many instances where subsynchronous vibrations from individual pads did not appear in any shaft data An important observation is that the shaft SSV hash indications most often correlated with one or more of the side pads, depending on bearing type and orientation The orientation and pad number schematic in Figure 7 is helpful in explaining the following observations and trends

In the case of load-between-pad orientation under some light downward loading, side pad 1, side pad 3, or both tended to correlate with the shaft SSV hash indications Upper pad 2 had more and higher subsynchronous vibration than the side pads in many cases, but rarely correlated with the shaft indications Noted earlier, shaft SSV hash decreased with applied load in the case of load-between-pad orientation, which is reasonable considering the high horizontal stiffness of this orientation with the shaft supported between the two bottom pads At the same time, pads 1, 2, and 3 become less strongly coupled to the shaft and may still experience subsynchronous vibrations, but with hydrodynamic forces too weak to affect the shaft

For a load-on-pad orientation under some light downward loading, the side pads (1, 2, 3, and 5), individually or in combinations, tended to correlate with the shaft SSV hash indications The correlation varied with bearing type and operating conditions, and

is also likely influenced by any slight skew in applied load vector and differences in manufacturing heights of the individual pads for this particular orientation Noted earlier, shaft SSV hash did not decrease as much with applied load for load-on-pad orientation, which is reasonable considering the weak horizontal stiffness for this orientation Pad 3 and pad 5 tended to correlate more with the shaft SSV hash indications as load was applied At the same time, pads 1 and 2 become less strongly coupled to the shaft and may still experience subsynchronous indications, but with hydrodynamic forces too weak to affect the shaft

SSV Grooves, Recent Tests

The SSV groove modification (Figure 5), developed during initial investigations in 1999, has since been successfully applied in

many compressor, turbine, and gearbox applications At the time,

development was based on hypotheses and shaft vibration data

Upgrades in the test rig capability and in high-speed data acquisition

equipment has more recently allowed an investigation of the

groove’s influence on pad vibration

SSV grooves were machined in the leading-edge-groove pads

used for results presented in Figure A-1 and were tested over the full

range of operating conditions for load-on-pad and load-between

pad orientation Test results for the same conditions and orientation

as Figure A-1 are presented in Figure A-2 (refer to APPENDIX A)

as an example comparison The tests found that subsynchronous

levels in all pads were significantly reduced or eliminated over the

full range of test operating conditions and configurations, and

confirmed the effectiveness of the method in eliminating shaft SSV

hash in evacuated, direct lube bearing designs

THEORETICAL INVESTIGATIONS

To help shed some light on these SSV hash measurements, the

bearing geometry and operating conditions corresponding with the

data in Figure A-1 were modeled using an algorithm developed by He

(2003) This algorithm includes the effects of supply flowrate in its

determination of a bearing’s steady-state (eccentricity, temperatures,

power loss, etc.) and dynamic (stiffness and damping) performance

based on the theories and experiments put forth by Heshmat

(1991) Since the measurements presented in these SSV hash tests

demonstrate the importance of supply flow on SSV, any theoretical

investigation must account for supply flowrate effects

Supply Flowrate Effects

The model applies to any bearing configuration where the

supply flowrate is not sufficient to ensure a full film across all the

pads’ surfaces In this case, the model will predict a partial

starvation at the inlet region of some of the pads Unlike the full

film situation where a continuous lubricant film starts at the pad

leading edge, in a partially starved situation, the continuous film is

formed downstream at some point where the clearance is sufficiently

reduced This physical phenomenon was observed and studied by

Heshmat (1991) on a two axial groove, sleeve bearing

Figure 22 illustrates predicted starvation effects for pad number 3,

corresponding with the test data pad numbers in Figure A-1 With a

supply flowrate of 8 gpm to the bearing, the film is able to generate

pressure across the pad’s entire arc length Reducing the bearing flow

to 4 gpm lowers overall pressure distribution, but the pad still

produces pressure across almost all of its arc length The pad is

clearly partially starved at 3 gpm where no film pressure is produced

at the inlet film region Decreasing the flow to 2 gpm expands the

starved inlet region At this flow level, pressure levels are very small

and the pivot is nearly centered within the positive pressure region

Partial starvation has effectively reduced the pad’s pivot offset, which

may explain unexpected results in center versus offset pivot tests

Figure 22 Partial Starvation Effects on Pad Film Pressure.

Pad Responsiveness

Since test results confirm that the individual pad motions are involved in this SSV hash phenomenon, the tilting-pad bearing’s full coefficients predicted by the modeling algorithm are of primary interest For a tilting-pad bearing with Npadpads, the full coefficients consist of 5Npad+4 stiffness and 5Npad+4 damping coefficients Described in detail by Shapiro and Colsher (1977) and Parsell, et al (1983), these full coefficients relate pad and shaft motions to forces and moments on the shaft and pads For rotordynamic purposes, the magnitude of each individual coefficient

is overlooked, focusing instead on their combined or “reduced” coefficients (four stiffnesses and four dampings) Closer scrutiny and understanding of these full coefficients are required when pad dynamics become the focus

Table 1 presents the stiffness full coefficients for a 60 percent offset, LOP, five-pad bearing with 12 gpm supply flowrate where:

⌬Fx and ⌬Fy = Horizontal and vertical forces on the shaft

⌬x and ⌬y = Horizontal and vertical shaft displacements

⌬M1… ⌬M5 = Moments on pads 1 through 5

⌬␦1… ⌬␦5 = Tilt angle changes on pads 1 through 5

Table 1 Stiffness Full Coefficients for the 60 Percent Offset LOP Bearing, 11,475 rpm, 200 psi.

Each value in Table 1 represents a stiffness, either ⌬Fx or ⌬M divided by ⌬x, ⌬y, or ⌬␦ depending on the position in the table Examining a few of these Table 1 coefficients will help explain the physical meaning behind them Tilt angle changes of the fourth pad (⌬␦4) produce the largest vertical force on the shaft (⌬Fy), since it has the highest value of Ky␦(⫺3.261e6 lbf/rad) This is due to the fourth pad being on the bottom, supporting the majority of the load Conversely, tilt angle changes on the second pad, located in the upper half of the bearing, produce the smallest vertical force on the shaft (Ky␦= 0.499e6 lbf/rad) The fourth pad is also the most difficult to tilt with the highest tilting stiffness K␦␦value (1.614e6 lbf-in/rad), while the second pad is the easiest to tilt (K␦␦= 0.245e6

lbf-in/rad) Tilting motions of the other four pads do not directly

cause any moments on the loaded pad The pads can only effectively communicate to each other through the shaft

Some of the full coefficients along with the pads’ polar inertia (Ip) form the equation of motion for pad tilting A simplified version of this equation, using a pad’s tilting stiffness K␦␦and damping C␦␦, is given by:

where Mextis an external moment being applied This equation can

be used to examine the frequency response characteristics of each pad’s tilting motion Each pad’s frequency response function (FRF) H(␻) is given by:

An FRF’s magnitude indicates how responsive a pad is to external moments trying to excite it Examining the predicted FRFs of each pad as a function of supply flowrate provides for a direct comparison with the pad waterfall diagrams of the test data

Figure 23 presents the predicted H(␻) of pad 3 in Figure A-1 as

an example, focusing on the low frequency region below 60 Hz

Regardless of flow, the frequency response remains relatively flat

as a function of excitation frequency with no peaks present This is

because the pad’s tilting motions remain overdamped for all the

flowrates examined An overdamped system (critical damping

greater than 100 percent) has no damped natural frequency,

resulting in no peaks in its response

Figure 23 H(␻) with Decreasing Supply Flow.

It is unclear as to the exact cause of the 8 Hz peak in Figure

A-1’s waterfall diagram for pad 3 One explanation is that there

may be a predominant excitation at this frequency Heshmat

(1991) observed a pulse excitation at similar low frequency

levels in his starvation experiments on a sleeve bearing

Another possibility may be that the pad is actually underdamped,

which the model has not been able to predict In this case, a

broadband excitation would result in a definitive peak at the

pad’s damped natural frequency Nothing indicates that the

peak is associated with an unstable, self-excited phenomenon,

since the pad’s tilting damping C␦␦ remains greater than or

equal to zero All indications are that the observed vibrations

are forced in nature

Both the measurements in Figure A-1 and the predicted FRF in

Figure 23 show that the third pad’s overall responsiveness does

dramatically increase as supply flow is reduced According to

the predictions in Figure 23, this pad responds very little to

excitations until the bearing flow is reduced below approximately

4 gpm This increased sensitivity correlates well with the predicted

pressure profiles in Figure 22 Below 4 gpm, the predictions in

Figure 22 show that the pad becomes partially starved at the

leading edge with low overall film pressures These lower film

pressures allow the pad to respond to external moment excitations

more easily

According to the H(␻) equation, the low frequency responsiveness

of an individual pad is largely dictated by its tilting stiffness, K␦␦

Figure 24 presents the calculated tilting stiffness of all five pads as

supply flow is varied With the exception of the fourth pad,

reducing supply flow causes a decrease in all the pads’ tilting

stiffnesses Below 5 gpm, the fourth pad’s stiffness begins to

increase as it supports more of the entire bearing load The strength

of this tilting stiffness means the fourth pad does not easily respond

to excitations, which correlates well with its relatively low vibrations

in Figure A-1

Figure 24 Tilting Stiffness of Individual Pads Versus Supply Flow.

When a pad becomes unloaded, it applies no pressure force on the shaft and its full coefficient stiffness and damping terms go to zero The pad then moves like a rigid body and its tilting FRF is dictated by the polar inertia The upper pads (1 and 2) become unloaded first as flow is reduced (Figure 24) This is because their larger film thicknesses require the most flow to maintain a full film along their entire surface With smaller film thicknesses, the third and fifth pads remain loaded until below 2 gpm These two pads’ tilting stiffness values are low in magnitude at reduced flowrates, resulting in the increased responsiveness observed in Figure 23 CONCLUSIONS

A series of tests and analyses were performed to investigate a peculiar, low-frequency, low amplitude, broadband subsynchronous

vibration, termed SSV hash, that has been witnessed in different

types of turbomachinery using tilting pad journal bearings Based on a study of test results for conventional and direct lube designs over a broad range of speeds, loads, and oil flows, the following shaft SSV hash trends were found to be fairly common for all test bearings:

• There were more SSV hash indications at low flow and low load, although amplitudes were sometimes higher at other operating conditions

• SSV hash levels at light loads were similar for load-between-pad and load-on-pad bearing orientations

• Shaft SSV hash decreased with applied load for load-between-pad orientation, but there was only a small change with load for load-on-pad orientation

• Consequently, load-on-pad operation produced higher SSV hash levels than load-between-pad orientation at higher loads

The following general trends for pad/shaft vibration correlation were found fairly common for all test bearings based on tests of bearing designs with vibration measurements of all five pads:

• All shaft SSV hash indications were confirmed to correlate with vibrations of at least one of the five pads

• The converse is not true There were subsynchronous vibrations in individual pads that did not appear in measured shaft vibration data

• The side pads most often correlated with shaft SSV hash indications The top upper pad for load-between-pad orientation had the highest subsychronous indications in many cases, but rarely correlated with any shaft data

The following specific observations were noted in test data

comparisons of bearing types and geometries:

• It is incorrect to associate SSV hash only with high-speed,

low-load applications Indications were noted over a broad range of

operating conditions

• Shaft SSV hash has been observed in conventional bearing designs

at low oil flow The data indicate that merely flooding the cavity is

insufficient to suppress shaft SSV hash At these conditions, direct

lube designs operate with lower levels of shaft SSV hash, attributed

to the more direct application of oil to the film

• At intermediate test oil flows, shaft SSV hash is more pronounced

in evacuated, direct lube bearings and can be reduced, but not

necessarily eliminated, by increasing oil flow Amplitudes appear

sensitive to certain operating conditions, around which are conditions

where the evacuated direct lube bearings operated with negligible

SSV hash indications

• There were mixed results in SSV hash comparisons of center

versus offset pivot direct lube test data The offset pivot pads in

most cases operated with the same to slightly less SSV hash

indications than the center pivot pads

• Comparisons among three direct lube methods noted distinct

differences in shaft SSV hash signatures, the results suggesting that

the differences originate at the entrance or leading edge of the

oil film

Overall, the various tests and comparisons indicate there are

factors other than flooding and bulk oil flow that contribute to SSV

hash behavior

There were two solutions determined from test data that

eliminated the SSV hash signature:

• The first required flooding the bearing cavity, using labyrinth or

floating seals and increased oil flow The solution was effective for

direct lube designs as well, but resulted in higher power loss, flow

requirements, and pad temperatures comparable to a conventional,

flooded design

• The second was a modification consisting of patented SSV

grooves cut in the babbitt near the edges of the pad, to capture and

redirect side leakage toward the leading edge of the next pad This

solution was successful in eliminating SSV hash in an evacuated

configuration Low oil flow and power loss were maintained, with

a slight penalty in pad temperature due to the introduction of warm,

side leakage oil back into the oil film

• In both cases, the elimination of SSV hash increased synchronous

amplitudes

The following conclusions are derived from the theoretical

investigation of SSV hash:

• A tilting pad bearing’s full coefficients can be used to assess the

dynamics of individual pads

• Using the full coefficients of one of the test bearings, theoretical

investigations suggest that the observed SSV hash is likely a forced

vibration phenomenon

• Theoretical predictions indicate that a pad’s responsiveness at

low frequencies increases when there is insufficient oil to provide

for a full film

• At lower supply flowrates, predicted partial starvation at the

leading edge progresses, which reduces the pad’s tilting stiffness,

making it more responsive to excitation

DISCUSSIONS

Many questions arose over the course of the tests and analyses

and during review of the initial drafts of this paper The following

discussions comment on topics that are not necessarily derived

directly from the test results and analyses As such, the discussions are subject to debate, which is welcome

Possible Source of Excitation

Test results and analyses indicate that the shaft SSV hash indications are caused by pad vibration with response characteristics indicative of a forced vibration Since there were no means in the SSV hash test apparatus to visualize the flow in the bearing, possible sources of excitation can only be surmised from the test results presented in this paper Instead, attention is directed to Heshmat (1991), who did observe a periodic phenomenon in his visualization experiments of oil streamlets in an axial groove journal bearing The phenomenon is reported as a “pulse” that periodically transformed the starved region of the leading edge film into a pattern of oil streamlets This pulse occurred every 0.5 to 1.0 seconds and its frequency varied with the degree

of starvation

This phenomenon is consistent with many of the observations from these SSV hash investigations Test data and analyses both show an increased sensitivity to SSV hash at reduced flowrates

A periodic pulse excitation can explain dominant SSV hash amplitudes at intermediate operating conditions, i.e., peaks noticeable

in Figures 17 through 22, as well as difference in SSV hash signatures among test bearings, which may be attributable to the influence of the different supply methods on the excitation

Pad Vibration and Damage

Pad flutter, babbitt fatigue, and pivot fretting are often brought

up as topics of concern regarding SSV hash The babbitt damage studied by Adams and Payandeh (1982) was a major concern in large, conventional, flooded bearings, attributed to self-excited, subsynchronous vibrations of unloaded pads with frequency ratios

approximately 0.50 that of running speed The term pad flutter is often used to describe this motion, and spragging is often used

to describe more violent, full clearance vibrations with forces sufficient to fatigue the babbitt and fret the pivots

The pad vibrations associated with SSV hash do not conform to these characteristics Subsynchronous amplitudes are at least an order of magnitude less than the bearing clearance, and frequency ratios are more on the order of 0.10 on average Babbitt fatigue is not considered a concern in regard to the SSV hash characteristics described in this paper

Fretting is a more difficult phenomenon to assess analytically Vibrations certainly contribute to fretting, but there are many sources of vibration in turbomachinery The only information that can be offered from the SSV hash test series in this respect is that there were no indications of fretting, or babbitt fatigue, in individual test pads after 600 hours of operation

Bearing Clearance, Preload, Pivot Offset

It would seem that preload, via reduced bearing clearance, would reduce shaft SSV hash and that increasing the pivot offset should make it worse It was therefore unexpected when center versus offset pivot tests showed lower levels for offset pivot operation Hindsight from the test results and analyses indicate other factors are involved, e.g., proximity of the individual pads to the shaft, the magnitude of their hydrodynamic forces, sensitivity to excitation, etc., which make it difficult to generalize cause and effect Although preload and clearance tests were not complete at the time of this paper, scenarios where preload might invoke pad vibration and shaft SSV hash can be envisioned when considering the other influences

With multiple factors and complex pad/shaft interactions, cause and effect scenarios for clearance, preload, offset, and other geometry should be performed by analysis rather than generalizations The analyses presented in the theory section of the paper are useful tools for assessing new designs, as well as evaluating changes to existing designs